Stationary gas turbines with a high power output have long been an essential component of power stations, especially combined-cycle power stations. FIG. 1 shows a perspective, partially sectional view of an example of such a gas turbine which is supplied by the Assignee of the present invention and is known by the type designation GT26®.
The gas turbine 10 of FIG. 1 is equipped with what is known as sequential combustion. It comprises a multistage compressor 12 which sucks in air via an air inlet 15 and compresses it. The compressed air is used, in a following first annular combustion chamber 14a, partially for the combustion of an injected fuel. The hot gas occurring flows through a first turbine 13a and then enters into a second combustion chamber 14b where the remaining air is employed for the combustion of a fuel which again is injected. The hot gas stream coming from the second combustion chamber 14b is expanded in a second turbine 13b so as to perform work and emerges from the gas turbine 10 through an exhaust gas outlet 16, in order to be discharged outward or, in a combined-cycle power station, in order to be used for the generation of steam.
The compressor 12 and the two turbines 13a, 13b have sets of moving blades which rotate about the axis 30 and which, together with guide vanes fastened to the surrounding stator, form the blading of the machine. All the moving blades are arranged on a common rotor 11 rotatable about the axis and are fastened releasably to the rotor shaft by means of rotor grooves provided for this purpose. Special attention is in this case devoted to the last stages 12a of the compressor 12 where the compressed air reaches temperatures of several hundred degrees Celsius.
It is known from the prior art (see, for example, WO-A1-2005/054682), according to FIG. 2, to provide the moving blades 12 of the last stages 12a of the compressor 12 with a blade root 18 designed as a hammerhead root and to push them with the blade root 18 into a rotor groove 19 extending about the axis and hold them there. The blade root 18 is supported on radial stop faces 25 of the rotor groove 19 which lie further outward in the radial direction, against centrifugal forces which act on the moving blade 17. Said blade root is likewise supported on axial stop faces 20 lying further inward in the radial direction, against axial forces which act on the moving blade 17. An undercut is in this case provided between each of the radial stop faces 25 and each of the axial stop faces 20. A spring 22 is provided at the bottom of the rotor groove 19 and fixes the moving blade 17 in the radial direction during assembly.
In the course of ongoing discussions about energy and the environment, there is the persistent desire to increase the power, efficiency, combustion temperature and/or mass throughflow of machines of this type. An increase in the power output can be achieved, inter alia, by improving the compressor.
An improvement in the gas turbine entails an increase in the mass throughflow through the compressor which leads to a higher gas temperature in the last compressor stages 12a. The up-to-date, progressive aerodynamic design of the blade leaves for the compressor requires greater axial chord lengths, this leading to a greater distance between the rotor grooves 19.
The two together give rise to markedly increased thermal stresses in the notches at the bottom of the rotor grooves in the rear compressor stages when the machine is being started, because the center of the rotor body is still at a low temperature (T1 in FIG. 2), whereas the outer region is already exposed to the high full-load temperature (T2 in FIG. 2), and therefore high thermal stresses occur in the material.
In another context, to be precise in moving blades of gas turbines with a dovetail-shaped blade root which bears against oblique stop faces in the rotor groove and because of the friction exerts shear forces on the side walls of the groove, it has been proposed to introduce fillets into the rotor groove below the stop faces in order to break down the friction-induced stresses (see U.S. Pat. No. 5,141,401). Here, however, thermal stresses do not play any part.
In connection with measures for reducing the stresses in the region of the rotor groove, EP-A1-1703080 repeats the critical influence of the cross-sectional contour of the groove upon the stress profile in the rotor. It is suggested there, in this connection, that the groove bottom be given an elliptical cross-sectional contour.
A rotor groove designed in this way has at its bottom, in order to reduce thermal stresses, an axially and radially widened bottom region 23 with a continuously curved cross-sectional contour which is distinguished by a large radius of curvature in the region of the mid-plane 33 and is designed to be mirror-symmetrical with respect to the mid-plane 33.
Should the design of the rotor root 18 of the moving blade 17 be preserved in the case of a rotor groove geometry modified in this way, the hammerhead of the blade root 18 according to FIG. 3 would have to be enlarged by the amount of the additional volume 24 illustrated by hatching, and this would lead to a marked increase in the mass of the moving blade 17 and therefore to a rise in the centrifugal forces acting on the rotor groove 21.